Centrifugal fluid machine

Abstract: At a vaned diffuser or a volute casing of a centrifugal fluid machine, pressure pulsation and vibrating forces acting upon the diffuser or the volute casing are mitigated or cancelled so as to abate the noise from the centrifugal fluid machine. The fluid machine having an impeller 3 rotating about a rotating shaft 2 within a casing 1 and having a vaned diffuser 4 or volute 12 fixed to the casing 1 is constructed such that radius of the vane trailing edge of the impeller 3 and radius of the vane leading edge of the diffuser 4 or radius of the volute tongue is varied in the direction of axis of rotation and inclinations, on a meridional plane, of the vane trailing edge of the impeller 3 and the vane leading edge of the diffuser 4 or the volute tongue are set in the same orientation, thereby reduction in head and efficiency or occurrence of an axial thrust may be restrained to the extent possible to optimally abate the noise and pressure pulsation of the centrifugal fluid machine.

Description: The present invention relates to centrifugal fluid assemblies such as a pump or compressor and, more particularly, relates to a centrifugal fluid assembly in which noise and pressure pulsation may be suitably abated. A flow distribution which is not uniform in the peripheral direction occurs at the outlet of an impeller due to the thickness of a vane and a secondary flow or boundary layer occurring between the vanes. Such nonuniform pulsating flow interferes with the leading edge of the vanes of a diffuser or a volute tongue, resulting in a periodical pressure pulsation and causing noise. In some cases, such pressure pulsation vibrates the diffuser and furthermore vibrates a casing or an outer casing outside thereof through a fitting portion, whereby the vibration is propagated into the air surrounding the pump to cause noise. Some proposals for reduction of pressure pulsation and noise in centrifugal assemblies are known from prior art. WO-A-93/10358 discloses a centrifugal compressor in which trailing edges of blades of a working wheel are provided with depressions diminishing the rotation radii of these edges into the body of the blades. That means that according to WO-A-93/10358 a radial distance between an axis of rotation and said trailing edge of the working wheel blade, measured along a perpendicular on said axis of rotation, is made smaller at the center of said working wheel blade trailing edge than at the two ends of said working wheel blade trailing edge. Stationary elements of an outlet system enter these depressions and have a form following the profile of the depressions. US-A-2 362 514 discloses a centrifugal compressor comprising a casing, an impeller located in the casing and having a plurality of circumferentially spaced blades. Furthermore a diffuser is located in the casing and surrounds the impeller for converting a part of the velocity energy of a medium discharged from the impeller into pressure energy. The diffuser has a plurality of circumferentially spaced vanes, whereby the diffuser vanes and the impeller blades have adjacent edges bevelled in opposite directions toward the axis of rotation. The diffuser vanes and the impeller blades comprise adjacent portions inclined in opposite directions with reference to planes through their roots and perpendicular to the plane of rotation. FR-A-352 787 discloses a diffuser type mixed flow pump, i.e. FR-A-352 787 is directed to a combination impeller/diffuser. With the arrangement as disclosed in FR-A-352 787 the flow has velocity components not only in a diametrical direction but also in an axial direction at the outlet of the impeller and at the inlet of the diffuser. In the arrangement as disclosed in FR-A-352 787, both the shroud and the hub are inclined in the same direction, and the flow passage defined by the shroud and the hub is inclined upwardly rightward. Thus, fluid flows upwardly rightward in the impeller and flows out at the outlet of the impeller in an upward and rightward direction. The same is the case with the stationary flow passage which is defined by the diffuser vanes and is formed to extend from the inlet to be directed upwardly rightward, the fluid flowing into the passage upwardly rightward. In the mixed flow pump as disclosed in FR-A-352 787 the trailing edge of the impeller vane and the leading edge of the diffuser vane are inclined in the same direction in configuration projected onto the meridional plane but both the impeller vane trailing edge and the diffuser vane leading edge are not offset relative to each other in a circumferential direction in front views. Accordingly, the fluctuating flow issuing from the impeller reaches the diffuser vane leading edge simultaneously over an area from the shroud side to the hub side, so that the fluctuating flow interferes much with the diffuser vane leading edge to generate much noise. US-A-3 628 881 discloses a scheme for reducing the amplitude of fluidborne noise produced by a centrifugal pump which comprises an improved impeller and in which the vanes are arranged in a single row and are skewed with respect to the shrouds so that the tips of adjacent vanes overlap in the circumferential direction. US-A-2 160 666 discloses a centrifugal-type fan with a scroll and a fan wheel consisting of a hub to which is secured a plurality of blades. The blades are provided with curved front ends. The curved front ends extend in direction of rotation of the fan wheel. In the vicinity of a point at which the blades are secured to the hub the blades are inclined rearwardly in axial direction from the direction of rotation from said point at which the blades are secured to the hub. A curved orifice member mounted in an intake opening of the scroll serves as a stationary part. A shroud ring formed as a substantial continuation of the orifice member is secured to the blades. In a centrifugal pump as disclosed in Sulzer Technical Review Vol.62 No.1 (1980) PP.24-26, the noise is reduced by varying the radius of the trailing edge of the vanes of the impeller or the peripheral position of the trailing edge of the vanes in the direction of the axis of rotation. Further, in an electric fan as disclosed in Japanese Patent Laid-Open Publication No.51-91006, a pressure increasing section and a noise abatement section (the noise abatement section being the portion where the peripheral position of a volute tongue is varied in the direction along the axis of rotation) are formed on the volute wall of a volute casing and the peripheral distance of the noise abatement section is made substantially equal to the peripheral distance between the trailing edges of the vanes that are next to each other in the impeller, so that the flow from the impeller does not impact the volute tongue all at once. In this manner, a shift in phase in the direction along the axis of rotation occurs in the interference between the flow and the volute tongue, whereby the periodical pressure pulsation is mitigated to lead to an abatement of the noise. In the above-described prior art, however, there has been a problem that, when the radius of the trailing edge of the vane of the impeller is varied in the direction along the axis of rotation, the head or the efficiency thereof is reduced due to the fact that the ratio between the radius of the trailing edge of the impeller vane and the radius of the leading edge of the diffuser vane or the radius of the volute tongue is varied in the direction along the axis of rotation. Further, when the outer radius of the main shroud and the front shroud of the impeller are different from each other in association with the fact that the trailing edge radius of the impeller vane is varied in the direction along the axis of rotation, an axial thrust occurs due to the difference between the projected areas of the main shroud and the front shroud in the direction along the axis of rotation. In the case where the peripheral position of the trailing edge of the impeller vane is varied in the direction along the axis of rotation, although the peripheral distance between the trailing edge of the impeller vane and the leading edge of the diffuser vane or the volute tongue is varied, the amount of such change has not been optimized. In the case where the peripheral position of the volute tongue is varied in the direction along the axis of rotation and the amount of such change is substantially equal to the peripheral distance between the trailing edges of the impeller vanes which are next to each other, the portion for effecting the pressure recovery in the volute casing becomes shorter whereby a sufficient pressure recovery cannot be obtained. An object of the present invention is to provide a centrifugal fluid assembly and a multistage centrifugal machine in which reduction in head and efficiency or occurrence of an axial thrust is controlled while noise and pressure pulsation are abated, and to indicate a preferred use of said centrifugal fluid assembly. According to the invention this object is achieved by a centrifugal fluid assembly according to claim 1. Furthermore the object is achieved according to the invention by a multistage centrifugal fluid machine according to claim 7. According to the invention the object is further achieved by indicating a preferred use of a centrifugal fluid assembly according to claim 3. Preferred embodiments of the centrifugal fluid assembly according to claim 1 are subject matter of claims 2 and 4 to 6. Preferred embodiments of the invention are described below with respect to the accompanying drawings in which: Fig.1 is a sectional perspective view of a diffuser pump showing an embodiment of the present invention. Fig.2 is a sectional view of a diffuser pump showing an embodiment of the present invention. Fig.3 is a detailed front sectional view taken along section III-III of Fig.2. Fig.4 is a development obtained by projecting the trailing edge of the impeller vane and the leading edge of the diffuser vane onto A-A circular cylindrical section of Fig.3. Fig.5 is a sectional view of a diffuser pump showing an embodiment of the present invention. Fig.6 is a sectional view of a diffuser pump showing an embodiment of the present invention. Fig.7 is a sectional view of a diffuser pump showing an embodiment of the present invention. Fig.8 is a sectional view of a diffuser pump showing an embodiment of the present invention. Fig.9 is a detailed front sectional view of a diffuser pump showing an embodiment of the present invention. Fig.10 is a sectional view of a diffuser pump showing an embodiment of the present invention. Fig.11 is a detailed front sectional view taken along section XIII-XIII of Fig.10 showing an embodiment of the present invention. Fig.12 is a development obtained by projecting the trailing edge of the impeller vane and the leading edge of the diffuser vane onto the A-A circular cylindrical section of Fig.11. Fig.13 is a development of another embodiment obtained by projecting the trailing edge of the impeller vane and the leading edge of the diffuser vane onto the A-A circular cylindrical section of Fig.11. Fig.14 is a sectional view of a barrel type multistage diffuser pump showing an embodiment of the present invention. Fig.15 is a sectional view of a horizontally split type multistage centrifugal compressor showing an embodiment of the present invention. Fig.16 is a barrel type single stage pump showing an embodiment of the present invention. Fig.17 is sectional view of a multistage mixed flow pump showing an embodiment of the present invention. Fig.18 illustrates flow distribution at the outlet of an impeller. Fig.19 shows frequency spectrum of the noise and pressure fluctuation of a pump. Fig.20 shows frequency spectrum of the noise and pressure fluctuation of a pump to which the present invention is applied. Fig.21 illustrates the direction along which the pressure difference force between the pressure surface and the suction surface of impeller vane is acted upon according to the present invention. An embodiment of the present invention will now be described by way of Fig.1. An impeller 3 is rotated about a rotating shaft 2 within a casing 1, and a diffuser 4 is fixed to the casing 1. The impeller 3 has a plurality of vanes 5 and the diffuser 4 has a plurality of vanes 6, where a trailing edge 7 of the vane 5 of the impeller 3 and a leading edge 8 of the vane 6 of the diffuser 4 are formed so that their radius is varied, respectively, along the axis of rotation. Fig.2 shows shapes on a meridional plane of a pair of impeller and diffuser as shown in Fig.1. The vane trailing edge 7 of the impeller 3 has its maximum radius at a side 7a toward a main shroud 9a and has its minimum radius at a side 7b toward a front shroud 9b. The vane leading edge 8 of the diffuser 4 is also inclined on the meridional plane in the same orientation as the vane trailing edge 7 of the impeller 3, and it has its maximum radius at a side 8a toward the main shroud 9a and its minimum radius at a side 8b toward the front shroud 9b. Fig.3 shows in detail the vicinity of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 of a section along line III-III of Fig.2. The impeller vane 5 and the diffuser vane 6 are of three-dimensional shape, i.e., the peripheral positions of the vanes are varied in the direction along the axis of rotation and radius of the impeller vane trailing edge 7 and radius of the diffuser vane leading edge 8 are varied in the direction along the axis of rotation, so as to vary the peripheral position of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 in the direction along the axis of rotation. The relative position in the peripheral direction between the impeller vane trailing edge 7 and the diffuser vane leading edge 8 of Fig.3 is shown in Fig.4. Fig.4 is obtained by projecting the impeller vane trailing edge 7 and the diffuser vane leading edge 8 onto a circular cylindrical development of the diffuser vane leading edge. In other words, of Fig.3, the impeller vane trailing edge 7 and the diffuser vane leading edge 8 as seen from the center of the rotating shaft are projected onto the cylindrical cross section A-A and it is developed into a plane. This is because in turbo fluid machines, a vane orientation is opposite between a rotating impeller and a stationary diffuser as viewed in a flow direction. By providing the inclinations, on a meridional plane, of the diffuser vane leading edge 8 and the impeller vane trailing edge 7 in the same orientation, a shift occurs in the peripheral position between the impeller vane trailing edge 7 and the diffuser vane leading edge 8. Due to such shift in the peripheral direction, the pulsating flow flowing out from the impeller vane trailing edge 7 impacts the diffuser vane leading edge 8 with a shift in phase so that the pressure pulsation is mitigated. Further, if the diffuser 4 is fixed to the casing 1 through a fitting portion 10 as shown in Fig.5, vibration of the diffuser 4 vibrated by the pressure pulsation propagates to the casing 1 through the fitting portion 10 and vibrates the surrounding air to cause noise; thus, the noise is abated when the pressure pulsation acting upon the diffuser vane leading edge 8 is mitigated according to the present embodiment. In the embodiment as shown in Fig.2, the shape of each of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 on a meridional plane is a straight line. In general, however, it suffices that the radius of the impeller vane trailing edge 7 and the radius of the diffuser vane leading edge 8 are monotonously increased in the direction along the axis of rotation, i.e. these radii are increased with the increase of the axial distance from the front shroud 9b, or monotonously decreased in the direction along the axis of rotation, i.e. these radii are decreased with the increase of the axial distance from the front shroud 9b, and inclinations of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 on a meridional plane are inclined in the same orientation, as shown in Fig. 6. Further, in the present embodiment shown in Fig.2, the outer diameters of the main shroud 9a and the front shroud 9b of the impeller 3 are, as shown in Fig.7, not required to be equal to each other and the inner diameters of the front shrouds 11a, 11b of the diffuser are not required to be equal to each other. By constructing in this manner, the ratio of the radii between the impeller vane trailing edge 7 and the diffuser vane leading edge 8 may be of the conventional construction, so that degradation in performance such as of head or efficiency due to an increase in the ratio of the radius of the diffuser vane leading edge to the radius of the impeller vane trailing edge does not occur. More preferably, as shown in Fig.8, by making the outer diameter of the main shroud 9a of the impeller 3 smaller than the outer diameter of the front shroud 9b, the vane length of the impeller may be made uniform from the main shroud 9a side to the front shroud 9b side, so that the projected area in the direction along the axis of rotation of the main shroud 9a on the high pressure side may be reduced with respect to the projected area of the front shroud 9b on the low pressure side so as to abate the axial thrust thereof. Further, as shown in Fig.3, the ratio (R a /r a ) of the radius R a of the outermost periphery portion 8a of the diffuser vane leading edge 8 to the radius r a of the outermost periphery portion 7a of the impeller vane trailing edge 7 is set the same as the ratio (R b /r b ) of the radius R b of the innermost periphery portion 8b of the diffuser vane leading edge 8 to the radius r b to the innermost periphery portion 7b of the impeller vane trailing edge 7, and the ratio of the radius of the impeller vane trailing edge to the radius of the diffuser vane leading edge is made constant in the axial direction, thereby degradation in performance may be controlled to a minimum. As shown in Figs. 2, 3, 5, 7 and 8, when the ratio between the trailing edge radius of the impeller and the leading edge radius of the diffuser vane is constant in the direction along the axis of rotation, pump performance is hard to exhibit drooping characteristics in a region of small flour rate. Further, Fig.9 illustrates in detail a case where the impeller vane 5 and the diffuser vane 6 are two-dimensionally designed. In Fig.9, vanes 5 and 6 are two-dimensionally shaped, i.e., the peripheral position of the vane is constant in the direction along the axis of rotation; however, by varying the radius of the impeller vane trailing edge 7 from the ontermost periphery portion 7a to the innermost periphery portion 7b and the radius of the diffuser vane leading edge 8 from the outermost periphery portion 8a to the innermost periphery portion 8b in the direction along the axis of rotation, the peripheral positions of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 are changed in the direction along the axis of rotation. For this reason, the pulsating flow impacts on the diffuser with a shift in phase so that force for vibrating the diffuser is reduced to abate the noise. Specifically, by forming the vanes into a two-dimensional shape, diffusion joining and forming of a press steel sheet thereof become easier and workability, precision and strength of the vane may be improved. The present invention as shown in Fig.2 or Fig.5 may be applied to a centrifugal pump or centrifugal compressor irrespective of whether it is of a single stage or of a multistage type. Another embodiment of the present invention will now be described by way of Fig.10. An impeller 3 is rotated about a rotating shaft 2 within a casing 1, and a diffuser 4 is fixed to the casing 1. The impeller 3 has a plurality of vanes 5 and the diffuser 4 has a plurality of vanes 6, where a trailing edge 7 of the vane 5 of the impeller 3 and a leading edge 8 of the vane 6 of the diffuser 4 are formed so that their radius is constant in the direction along the axis of rotation. Fig.11 shows in detail the vicinity of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 along cross section XIII-XIII of Fig.10. The impeller vane 5 and the diffuser vane 6 are of three-dimensional shape, i.e., the peripheral position of the vanes is varied in the direction along the axis of rotation. The relative position in the peripheral direction of the impeller vane trailing edge 7 and the diffuser vane leading edge 8 of Fig.11 is shown in Fig.12. Fig.12 is obtained by projecting the impeller vane trailing edge 7 and the diffuser vane leading edge 8 onto a circular cylindrical development of the diffuser vane leading edge. In other words, the impeller vane trailing edge 7 and the diffuser vane leading edge 8 as seen from the center of the rotating shaft in Fig.11 are projected onto the circular cylindrical section A-A and it is developed into a plane. As shown in Fig.12, the difference (ℓ 1 -ℓ 2 ) between the maximum value ℓ 1 and the minimum value ℓ 2 of the peripheral distance between the impeller vane trailing edge 7 and the diffuser vane leading edge 8 is made equal to the peripheral distance ℓ 3 between the vane trailing edges that are next to each other in the impeller. Since a pulsating flow of one wavelength occurs between the vane trailing edges that are next to each other in an impeller, the phase of the pulsating flow impacting the diffuser vane leading edge 8 is shifted exactly corresponding to one wavelength along the axis of rotation; therefore, pressure pulsation applied on the diffuser vane leading edge 8 due to the pulsation and the vibrating force resulting therefrom are cancelled when integrated in the axial direction. The present invention as shown in Fig.11 may be applied to a centrifugal pump or centrifugal compressor irrespective of whether it is of a single stage or of multistage type. Alternatively, by setting (ℓ 1 -ℓ 2 ) to a part obtained by dividing ℓ 3 into "n" (integer) identical parts, the phase of the pulsation flow impacting the diffuser vane leading edge 8 is shifted exactly corresponding to one wavelength of "n"th higher harmonic in the axial direction so that the vibrating forces acting on the diffuser vane leading edge 8 due to the "n"th higher harmonic component of fluctuation are cancelled when integrated in the axial direction. Especially in a multistage fluid machine or a fluid machine having armoured type casing, vibration is transmitted through a fitting portion between the stages or between the inner and outer casings so that the vibrating force due to the first or "n"th dominant frequency of the above pressure pulsation largely contributes to the noise; therefore, it is important for abating the noise to design so that, of the vibrating forces due to pulsating flow, specific high order frequency components contributing to the noise are cancelled. Furthermore, as shown in Fig.13 where the diffuser vane leading edge and the impeller vane trailing edge are projected onto a circular cylindrical development of the diffuser vane leading edge, by setting the impeller vane trailing edge 7 and the diffuser vane leading edge 8 perpendicular to each other on the circular cylindrical development, the direction of the force due to the pressure difference between the pressure surface and the suction surface of the impeller vane becomes parallel to the diffuser vane leading edge, whereby the vibrating force due to such pressure difference does not act upon the diffuser vane and the noise may be abated. The frequency spectrum of the noise and of pressure fluctuation at the diffuser inlet is shown in Fig.20 of the case where the embodiment shown in Fig.15 is applied to a centrifugal pump. This pump has a combination of such number of vanes that the vibrating frequencies of 4NZ and 5NZ are dominant; in the case of a conventional pump shown in Fig.19, the noise, too, is dominant at the frequency components of 4NZ, 5NZ. In the pump to which the present invention is applied, the dominance of 4NZ, 5NZ frequency components is eliminated with respect to the pressure fluctuation as shown in Fig.20, and, as a result, 4NZ, 5NZ frequency components are remarkably reduced also in the noise so as to greatly abate the noise. The invention shown by way of the embodiment of Fig.13 may be applied to abate the noise in a single stage or multistage centrifugal pump or centrifugal compressor having a fitting portion between the diffuser portion and the casing or between the inner casing and the outer casing. It should be noted that the embodiments of Fig.12 and Fig.13 may be achieved also by varying the radius of the impeller vane trailing edge and radius of the diffuser vane leading edge in the direction along the axis of rotation as shown in Fig.2. In other words, these correspond to special cases of the embodiment shown in Fig.4. The present invention as described above may be applied to a fluid machine having an impeller rotating about an axis of rotation within a casing and a vaned diffuser fixed to the casing; Fig.14 being an embodiment applied to a barrel type multistage diffuser pump; Fig.15 being an embodiment applied to a horizontally split type multistage centrifugal compressor; and Fig.16 being an embodiment applied to a barrel type single stage pump. Further, the present invention may be applied not only to centrifugal types but also to mixed flow types. Fig.17 shows an embodiment applied to a multistage mixed flow pump. Furthermore, the case where multistage fluid machines are used, it is important to know how to set the inclination on a meridional plane of the impeller trailing edge 7 for each stage. The reason for this is that: when, the outer radius of the main shroud 9a and the front shroud 9b of the impeller and the inner radius of the front shrouds 11a, 11b of the diffuser are different, respectively, while the radius ratio of the impeller and the diffuser may be smaller to control degradation in performance, the projected areas in the direction along the axis of rotation of the two front shrouds are different from the conventional art and there is a problem of axial thrust due to the difference in these areas. In the embodiment of Fig.14, the outer radius of the main shroud 9a of the impeller at all stages is smaller than the outer radius of the front shroud 9b. In this manner, the vane length of the impeller is made uniform from the main shroud 9a side toward the front shroud 9b, and the projected area in the direction along the axis of rotation of the main shroud 9a on the high pressure side may be made smaller in relation to the projected area of the front shroud 9b on the low pressure side, to thereby abate the axial thrust. In the embodiment of Fig.15, the inclination on a meridional plane of the impeller vane trailing edge is reversed between the stages that are next to the each other so that an axial thrust due to difference in the projected areas of the main shroud and the front shroud may be cancelled. Operation of the above described embodiments will now be described in further detail. A flow W 2 at the outlet of the impeller forms a flow distribution that is nonuniform in the peripheral direction as shown in Fig.18 due to the thickness of the vane 5, and the secondary flow and boundary layer between the vanes. Such nonuniform pulsating flow is interfered with a diffuser vane leading edge to generate periodical pressure pulsation which causes a noise. In other cases, such pressure pulsation vibrates the diffuser and furthermore vibrates a casing or an outer casing outside thereof through a fitting portion so that the vibration is propagated into the air surrounding the pump to cause noise. Frequency spectrum of the noise and of pressure pulsation at the diffuser inlet of a centrifugal pump is shown in Fig.19. The frequency of the pulsating flow is the product NxZ of a rotating speed N of the impeller and number Z of the impeller vanes, the frequency on the horizontal axis being made non-dimensional by NxZ. The pressure pulsation is dominant not only at the fundamental frequency component of NxZ but also at higher harmonic components thereof. This is because the flow distribution at the impeller outlet is not of a sine wave but is strained. The noise is dominant at specific higher harmonic components of the fundamental frequency component of NxZ and the noise is not necessarily dominant at all the dominant frequency components of the above pressure pulsation. This is because, as disclosed in Japanese Patent Unexamined Publication No.60-50299, when the pulsating flow is vibrating the diffuser vane, there are some frequency components for which the vibrating force is cancelled as to the entire diffuser and some other components for which it is not cancelled, due to the combination of the number of vanes of the impeller and the diffuser. Especially, the vibration is transmitted through a fitting portion between the stages or between the inner and outer casings in a multistage fluid machine or armoured type casing fluid machine, or, in the case of a single stage, between the diffuser and the casing, so that the vibrating force due to the above dominant frequencies largely contributes to the noise. The centrifugal pump of which the measured result is shown in Fig.19 is constituted by a combination of the number of vanes for which the vibrating frequencies are dominant at 4NZ and 5NZ, the noise being dominant also at the frequency components of 4NZ, 5NZ. Specifically, the vibrating force is increased as the nonuniform pulsating flow impacts the respective position in the direction along the axis of rotation of the diffuser vane leading edge with an identical phase. Accordingly, the pressure pulsation and the vibrating force may be reduced to abate the noise by shifting the phase of the pulsating flow reaching the diffuser vane leading edge, by forming an inclination on the diffuser vane leading edge or by forming an inclination on the impeller vane trailing edge. As shown in a meridional sectional view of Fig.2 and a front view of Fig.9 illustrating the impeller and the diffuser of a diffuser pump, the radius of the impeller vane trailing edge 7, and the radius of the diffuser vane leading edge 8 are varied in the direction along the axis of rotation; thereby the peripheral positions of the impeller vane trailing edge, and the diffuser vane leading edge are varied in the direction along the axis of rotation. In particular, in turbo fluid machines, a vane orientation is made opposite between a rotating impeller and a stationary diffuser as viewed in a flow direction. Accordingly, as shown in Fig.2, the radius of the impeller vane trailing edge, and diffuser vane leading edge is monotonously increased or decreased in the direction along axis of rotation and the impeller vane trailing edge, and the diffuser vane leading edge and are inclined in the same orientation on a meridional plane; thereby, as shown in Figs.4 and 12 where the impeller vane trailing edge and the diffuser vane leading edge are projected onto a circular cylindrical development of the diffuser leading edge portion, a shift occurs in the peripheral position between the impeller vane trailing edge 7 and the diffuser vane leading edge 8. Accordingly, the peripheral distance between the impeller vane trailing edge and the diffuser vane leading edge is varied in the axial direction, whereby the fluctuating flow flowing out from the impeller vane trailing edge impacts the diffuser vane leading edge with a shift in phase so as to cancel the pressure pulsation. For this reason, the vibrating force acting upon the casing is reduced and the noise is also abated. It should be noted that the change in the direction along the axis of rotation of the radius of the impeller vane trailing edge, and the radius of the diffuser vane leading edge is not limited to monotonous increase or decrease, and similar noise abating effect may be obtained by changing them in different ways. The present invention may be applied to the case where the diffuser vane, and the impeller vane are of two-dimensional shape, i.e., are designed so that the peripheral position of the vane is constant in the direction of axis of rotation (Fig.9) and to the case where they are formed into a three-dimensional shape, i.e., are designed so that the peripheral position of the vane is varied in the direction of axis of rotation (Fig.3). Especially, since abating of noise is possible with vanes having a two-dimensional shape, diffusion joining and forming of a press steel sheet are easier and manufacturing precision of the vanes may be improved. Further, since the inclinations on a meridional plane are in the same orientation, the ratio of the radius of the impeller vane trailing edge to the radius of diffuser vane leading edge the is not largely varied in the direction of axis of rotation whereby degradation in performance is small. In other words, pressure loss due to an increased radius ratio may be reduced to control degradation in head and efficiency. Further, by setting constant the ratio of the radius of the impeller vane trailing edge to the radius of the diffuser vane leading edge in the direction along the axis of rotation, degradation in performance may be controlled to the minimum. Other effects of the present invention will now be described by way of Fig.12. In Fig.12, the impeller vane trailing edge 7 and the diffuser vane leading edge 8 as seen from the center of the rotating axis in the front sectional view (Fig.11) of the impeller and the diffuser are projected onto a circular cylindrical section A-A and are developed into a plane. The peripheral distance between the impeller vane trailing edge 7 and the diffuser vane leading edge 8 is varied in the direction along the axis of rotation such that the difference (ℓ 1 -ℓ 2 ) between the maximum value ℓ 1 and the minimum value ℓ 2 of the peripheral distance between the impeller vane trailing edge and the diffuser vane leading edge is identical to the peripheral distance ℓ 3 between the vane trailing edges that are next to each other in the impeller. Since a pulsating flow corresponding to one wavelength is generated between the vane trailing edges that are next to each other in the impeller, the phase of the pulsating flow impacting the diffuser vane leading edge is shifted exactly by one wave length so that the pressure pulsation and vibrating force acting upon the diffuser vane leading edge due to the pulsation are cancelled when integrated in the direction along the axis of rotation. However, a rather large inclination is necessary to make the above difference (ℓ 1 -ℓ 2 ) equal to the peripheral distance ℓ 3 between the vane trailing edges that are next to each other in the impeller. As described above, when the pulsating flow at the outlet of the impeller vibrates the diffuser vane leading edge, only specific higher harmonic components of NZ frequency components are dominant and contribute to vibrating of the diffuser, depending on the combination of the number of impeller vanes and the number of diffuser vanes. Therefore, if the difference (ℓ 1 -ℓ 2 ) between the maximum value ℓ 1 and the minimum value ℓ 2 of the peripheral distance between the impeller vane trailing edge and the diffuser vane leading edge is made equal to one of equally divided "n" (integer) parts of the peripheral distance ℓ 3 between the vane trailing edges that are next to each other in the impeller, the phase of the pulsating flow impacting the diffuser vane leading edge is shifted exactly corresponding to one wavelength of "n"th higher harmonic in the direction along the axis of rotation so that the vibrating forces applied on the diffuser vane leading edge due to the "n"th higher harmonic component of the pulsation are cancelled when integrated in the direction along the axis of rotation. Especially in a multistage fluid machine or an armoured type casing fluid machine, vibration is transmitted through a fitting portion between the stages or between outer and inner casings whereby vibrating forces due to the above dominant frequencies largely contribute to the noise; therefore, it is important for abatement of the noise to design in such a manner that, of the vibrating forces due to the pulsating flow, specific high order frequency components contributing to the noise are cancelled. The above effect may also be obtained such that the impeller vane trailing edge and the diffuser vane leading edge are formed into a three-dimensional shape and, as shown in Fig.11, while the respective radius of the impeller vane trailing edge and the diffuser vane leading edge is fixed in the direction along the axis of rotation, only their peripheral positions are changed. In other words, if the difference (ℓ 1 -ℓ 2 ) between the maximum value ℓ 1 and the minimum value ℓ 2 of the peripheral distance between the impeller vane trailing edge and the diffuser vane leading edge is made equal to the peripheral distance ℓ 3 between the vane trailing edges that are next to each other in the impeller or to a part of "n" (integer) equally divided parts thereof, the first order or "n"th order vibrating forces applied on the diffuser vane leading edge is cancelled when integrated in the axial direction. Furthermore, when the diffuser vane leading edge and the impeller vane trailing edge are projected onto a circular cylindrical development of the diffuser vane leading edge, by setting the vane leading edge and the vane trailing edge perpendicular to each other on the above circular cylindrical development, it is possible to abate the vibrating force due to pressure pulsation applied on the diffuser vane leading edge. Thus, if, as shown in Fig.21, the impeller vane trailing edge and the diffuser vane leading edge are set perpendicular to each other, the direction of force F due to the pressure difference between the pressure surface p and the suction surface s of the impeller vane becomes parallel to the diffuser vane leading edge so that the vibrating force does not act upon the diffuser vane. In the case where, as shown in Fig.7, the outer diameter of the main shroud 9a of the impeller is made larger than the outer diameter of the front shroud 9b and the inner diameters of the two corresponding front shrouds of the diffuser are varied respectively in accordance with the outer diameters of the main shroud and the front shroud of the impeller, while the radius ratio of the impeller to the diffuser may be made smaller to control degradation in performance, a problem of an axial thrust occurs due to the fact that the projected areas in the direction along the axis of rotation of the main shroud and the front shroud are different from each other. Therefore, in the case of having a multiple of stages, in addition to varying the radius of the impeller vane trailing edge in the direction along the axis of rotation, the outer diameters of the main shroud and the front shroud are made different for at least two impellers; and, of those impellers for which the outer diameters of the main shroud and the front shroud are made different from each other, the outer diameter of the main shroud is made larger than the outer diameter of the front shroud for at least one impeller and the outer diameter of the main shroud is made smaller than the outer diameter of the front shroud for the remaining impellers; thereby, it is possible to reduce the axial thrust occurring due to the difference in the projected area in the direction along the axis of rotation of the main shroud and the front shroud. As has been described, according to the present invention, noise and pressure pulsation of a centrifugal fluid machine may be optimally abated with restraining to the extent possible degradation in head and efficiency or occurrence of an axial thrust.

Claims: A centrifugal fluid assembly comprising an impeller (3), rotating together with a rotating shaft (2) within a casing (1) about an axis of rotation and comprising at least one impeller vane (5) with an impeller vane trailing edge (7), and a diffuser (4) fixed to said casing (1) and comprising at least one diffuser vane (6) with a diffuser vane leading edge (8), wherein the radial distance between said axis of rotation and said impeller vane trailing edge (7), measured perpendicular to said axis of rotation, and the radial distance between said axis of rotation and said diffuser vane leading edge (8), measured perpendicular to said axis of rotation, both either monotonously increase with increasing axial distance or monotonously decrease with increasing axial distance over their whole axial extension, but not including the case where any of said radial distances is constant over the whole axial extension, projections of the impeller vane trailing edge (7) and of the diffuser vane leading edge (8) onto a meridional plane have the same orientation, and a shift occurs in the peripheral position between the impeller vane trailing edge (7) and the diffuser vane leading edge (8), due to the fact that the projections of these edges onto a circular cylinder (A-A: Figs. 3 and 11) coaxial to said axis of rotation are inclined in opposite directions with respect to said axis of rotations (Figs. 3, 4, 11, 12, 13, 21), and the radial distance between the projections of said impeller vane trailing edge (7) and of said diffuser vane leading edge (8) is constant in the axial direction. The centrifugal fluid assembly according to claim 1, characterized in that said impeller vane trailing edge (7) or said diffuser vane leading edge (8) or both have two-dimensional shape. Use of a centrifugal fluid assembly according to claim 1 in a barrel type centrifugal fluid machine comprising said casing as an inner casing (1) within an outer casing (1b). The centrifugal fluid assembly according to claim 1, characterized in that the difference (1 1 -1 2 ) between the maximum value (l 1 ) and the minimum value (l 2 ) of the peripheral distance between the impeller vane trailing edge (7) and the diffuser vane leading edge (8) is equal to the peripheral distance (l 3 ) between the impeller vane trailing edge (7) and the adjacent impeller vane trailing edge (7). The centrifugal fluid assembly according to claim 1, characterized in that it comprises a plurality of impeller vanes (5), the difference (l 1 -l 2 ) between the maximum value (l 1 ) and the minimum value (l 2 ) of the peripheral distance between said impeller vane trailing edge (7) and said diffuser vane leading edge (8) is equal to the peripheral distance (l 3 ) between said impeller vane trailing edge (7) and the adjacent impeller vane trailing edge (7) or to one of n equal parts of the peripheral distance (l 3 ) between said impeller vane trailing edge (7) and said adjacent impeller vane trailing edge (7), whereby n is an integer greater than 1. The centrifugal fluid assembly according to claim 1, characterized in that when said diffuser vane leading edge (8) and said impeller vane trailing edge (7) are projected onto said circular cylinder, the projections of the impeller vane trailing edge (7) and of the diffuser vane leading edge (8) are perpendicular to each other. A multistage centrifugal fluid machine comprising at least two centrifugal fluid assemblies according to claim 1 or at least two centrifugal fluid assemblies according to claim 4 or, at least two centrifugal fluid assemblies according to claim 6, characterized in that in an even number of said centrifugal fluid assemblies an impeller main shroud (9a) and an impeller front shroud (9b) have different outer shapes, wherein the outside diameter of the impeller main shroud (9a) of one half of said even number of said centrifugal fluid assemblies is made larger than the outside diameter of the corresponding impeller front shroud (9b), and the outside diameter of the impeller main shroud (9a) of the other half of said even number of said centrifugal fluid assemblies is made smaller than the outside diameter of the corresponding impeller front shroud (9b).

Inventors: IDA MICHIAKI, ISHIMARU HIROTOSHI, IWASAKI SABURO, IWASE YUKIJI, NAGAOKA YOSHIHIRO, TANAKA SADASHI, UEYAMA YOSHIHARU, YOSHIDA TETUYA, IDA, MICHIAKI, ISHIMARU, HIROTOSHI, IWASAKI, SABURO, IWASE, YUKIJI, NAGAOKA, YOSHIHIRO, TANAKA, SADASHI, UEYAMA, YOSHIHARU, YOSHIDA, TETUYA
Applicants: HITACHI LTD, HITACHI, LTD.
Application year: 1994
Main classification: F04D29
IPC codes: F04D29